Turbo-charged internal combustion engine with in-cylinder egr and injection rate shaping

ABSTRACT

A turbo-charged internal combustion cylinder assembly includes a combustion chamber which may be communicably connected to a compressor via an intake port through an intake manifold and aftercooler so the compressor may provide pre-combustion gases to the combustion chamber when the intake valve is open. An exhaust port communicably connects the combustion chamber to an exhaust manifold. An exhaust valve may open to exhaust post-combustion gases to the exhaust manifold while an intake valve is substantially closed, and the exhaust valve may open to admit post-combustion gases to the combustion chamber while the intake valve is substantially open and an exhaust port pressure in the exhaust port is higher than a combustion chamber pressure in the combustion chamber. A fuel injector may admit fuel to the combustion chamber. A spill valve may control a rate of fuel injection to the combustion chamber, the spill valve having a first position providing a maximum fuel injection rate, a second position providing a substantially zero fuel injection rate, and at least one intermediate position providing an intermediate fuel injection rate between the maximum fuel injection rate and the zero fuel injection rate.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a divisional of U.S. application Ser. No.10/372,929, filed Feb. 26, 2003, which claims the benefit of ProvisionalApplication Ser. No. 60/360,005, filed Feb. 28, 2002, the disclosure ofwhich is incorporated by reference.

FIELD OF THE INVENTION

The invention relates to internal combustion engines, and in particular,to turbo-charged internal combustion engines with fuel injection rateshaping and internal exhaust gas recirculation.

DESCRIPTION OF THE RELATED ART

Emission control standards for internal combustion engines have tendedto become more stringent over time. The sorts of emissions to becontrolled tend to fall into at least four broad categories: unburnedhydrocarbons, carbon monoxide, particulates, and oxides of nitrogen(NOx). Unburned hydrocarbons and carbon monoxide tend to be produced byinefficient or incomplete combustion. Efficient, complete combustion, onthe other hand, tends to produce oxides of nitrogen.

Efficient, complete combustion tends to be characterized by highcombustion chamber temperatures. The heat associated with highcombustion chamber temperatures acts as a catalyst, promoting thebinding of oxygen in the air charge to the otherwise inert nitrogen andproducing oxides of nitrogen. An engine that is running efficiently,therefore, may produce oxides of nitrogen. Controlling the amounts ofemissions produced by an internal combustion engine, then, becomes anissue of balancing combustion efficiency against raising combustiontemperatures high enough to produce oxides of nitrogen.

Since the ingredients of oxides of nitrogen come from the intake air,one possibility to reduce the amounts of oxides of nitrogen may be tolimit the air available for combustion. Compression-ignition engines,unlike spark-ignition engines, are often run with an excess of air overthe stoichiometric ratio, so there is lots of nitrogen available foroxidation. This is because the production of particulates, such as ash,tends to rise as the air/fuel (A/F) mixture approaches stoichiometric.This is evidenced by the observation that diesel trucks often emit puffsof smoke under heavy acceleration. Since compression-ignition enginesneed to run with an excess of air to avoid emitting particulates,reducing the production of oxides of nitrogen by reducing the amount ofair available for combustion is not a practical solution.

Another way to control the production of oxides of nitrogen is to reducepeak combustion chamber temperatures. Since production of oxides ofnitrogen tends to depend on high combustion chamber temperatures as acatalyst, reducing the peak temperature ameliorates one of theconditions necessary for the production of oxides of nitrogen. Reducingthe peak combustion chamber temperature may thus reduce the amount ofoxygen that binds with nitrogen, with a consequent reduction in thequantity of oxides of nitrogen produced.

One means of lowering the high combustion chamber temperatures producedby an efficient combustion event is to cool the combustion chamberduring combustion. The combustion chamber may be cooled by, e.g.reintroducing some of the products of previous combustion events backinto the combustion chamber, a process known as exhaust gasrecirculation (EGR). Since the products of efficient combustion areprimarily water and carbon dioxide, neither of which is very flammable,this has the effect of extinguishing the combustion somewhat. The peaktemperatures reached in the combustion chamber will consequently belower, which retards the production of oxides of nitrogen. Clearly, theamount and timing of the introduction of products of combustion must becontrolled accurately to avoid impairing the performance of the engine.

Lowering combustion chamber temperatures may have the collateral benefitof reducing exhaust manifold temperatures, as well as stacktemperatures. Reducing stack temperatures reduces the temperature inexhaust after-treatment equipment such as oxidation catalysts, with aconsequent reduction in the formation of, e.g. sulfates. Reducing stacktemperatures may also reduce the production of particulates.

One way to reintroduce some of the products of previous combustionevents into the combustion chamber is with external EGR. In externalEGR, a tube or plenum conducts some post-combustion gases from thecombustion chamber, usually through an exhaust manifold, to a valve.When the valve is opened, the post-combustion gases are readmitted tothe combustion chamber, often passing through the intake manifold first.If the post-combustion gases pass through the intake manifold they willmix with fresh make-up air coming in through the air cleaner and bedistributed relatively evenly to each of the combustion chambers whenits respective intake valve opens.

External EGR, however, relies on high engine heat rejection to work,since the post-combustion gases must travel a relatively long way. Also,the valves and other hardware associated with external EGR increase thecost and complexity of the engine. Furthermore, the addition of externalEGR and its associated hardware to an existing engine may require thechassis, front clip, or sheet metal to be re-arranged to allow theengine to fit. Furthermore, if the external EGR is plumbed through theintake manifold, it may be difficult to control the amount of exhaustgas that is re-admitted to each individual combustion chamber. This maypose a problem if, e.g. some combustion chambers run hotter than othercombustion chambers, such as those that are nearer the water jacketexit.

A combustion chamber near the exit to the water jacket will betransferring heat to warmer coolant, other things being equal, than acombustion chamber near, e.g. the entrance to the water jacket, sincethe coolant has already been past the other combustion chambers when itreaches the exit. There will thus be a smaller temperature differentialbetween the combustion chamber and the coolant. Thus the metal around,e.g. the combustion chamber will be maintained at a higher temperature,other things being equal. It would be desirable if the amount of exhaustgas that is readmitted to a combustion chamber could be controlled on anindividual basis, commensurate with the temperatures prevailing in thatcombustion chamber.

The EGR valve, along with the associated actuator and control hardware,is also a point of potential failure, jeopardizing the durability of theengine. It would be desirable if the EGR valve, and its associatedactuator and control hardware, could be eliminated. It would further bedesirable if the amount and timing of post-combustion gases thatre-enter the combustion chamber could be controlled by varying thepressure in the exhaust manifold relative to the pressure in thecombustion chamber, rather than with an external valve. Finally,allowing post-combustion gases to re-enter the combustion chamberdirectly from the exhaust manifold may reduce the transfer time of thepost-combustion gases back into the combustion chamber, improving theresponsiveness of the EGR system and allowing their application to beoptimized or, at least, reduced.

Many truck engines are supercharged. Some superchargers are belt-,chain- or gear-driven, while others, so-called turbo-chargers, rely on aturbine to convert the kinetic energy in exhaust gases to rotationalmomentum in a compressor. There are those who define superchargers andturbo-chargers as separate entities. For the purposes of thisapplication, however, a turbo-charger will be defined as aturbine-driven supercharger.

Turbo-machinery, such as superchargers, have components that rotate.These components possess inertia. These components gain momentum withrespect to this inertia when they are turned, by, e.g. a belt or aturbine. Building rotational momentum requires time, which manifestsitself as lag. The lag is generally proportional to the inertia of theturbine rotor and compressor. Thus the inertia of the turbine rotor andthe compressor rotor contribute to lag. There are advantages to be foundwith using smaller compressors and turbine trim, such as bettertransient engine response, which in turn helps to control emissions,such as, for example, particulates. It would be desirable to be able toreduce the sizes of the turbine and the compressor rotors, thus reducingthe lag normally associated with, e.g. turbo-chargers, and improving thetransient response.

Turbo-chargers rely on post-combustion gases for their energy.Sometimes, under operating conditions such as at start-up or low-speedoperation, an engine does not produce enough post-combustion gases todrive the compressor adequately. It would be desirable if a compressorhad a secondary source of power, such as an electrical or beltdriven-clutch-assist, for situations when more turbo boost is called forthan the available exhaust gas can produce. This-would be especiallydesirable if the turbo-charger were part of the emission control system.

Carnot taught that there are two ways to increase the efficiency of aheat engine, by raising the temperature at which heat is added or byreducing the temperature at which heat is rejected. Although every pointwithin a diesel engine combustion chamber should be at or above thekindling temperature of the fuel when fuel is admitted to the combustionchamber, this may not always be the case. The fuel itself may, e.g. becold relative to the combustion chamber, especially during winterdriving. Cold fuel may thus reduce the temperature locally in thecombustion chamber below the kindling temperature of the fuel.

Transient conditions such as those due, e.g. to start up or rapidchanges in throttle position may contribute to cooler combustion chambertemperatures as well. Throttling is a cooling process, and so fuel thathas been throttled will generally be cooled somewhat. It would bedesirable if the fuel being injected were pre-heated slightly by, e.g.using the heat of the post-combustion gases, so that it would be morelikely to be ignited completely upon entry into the combustion chamber.

Fuel is normally injected, on the average, into the center of acombustion chamber. Although average combustion chamber temperatures maybe relatively constant, local temperatures may fluctuate. Combustionchamber temperatures, for example, may vary both spatially across thecombustion chamber, and over time during the combustion event.

Since some points within a combustion chamber are hotter than others, itwould be desirable to be able to adjust the rate at which fuel isinjected. Thus, the rate at which fuel was injected could be varied atdifferent times and at different points within the combustion chamber,during the combustion event, so fuel was injected where and when thecombustion chamber temperatures are highest. This may, for example,allow the combustion process to rely less on propagation of a flamefront to burn the fuel. It may also allow the peak temperature to bereduced, thereby reducing formation of oxides of nitrogen, since thefuel can be directed at a point where the temperatures are highest.

Adjusting the rate at which fuel is injected is often termed injectionrate shaping. One means of injection rate shaping is described in U.S.Pat. No. 6,336,444 B1 to Suder, the disclosure of which is incorporatedby reference. It would be desirable for injection rate shaping to becombined with, e.g, a lash adjustment mechanism, improved turbo-chargerefficiency, or in-cylinder exhaust gas recirculation with and withoutpost bump shutoff capability.

SUMMARY OF THE INVENTION

In several aspects, the invention may provide post bump shutoff with alash adjustment mechanism, improved turbo-charger efficiency, a modifiedpost bump injection system, and in-cylinder exhaust gas recirculationwith and without post bump shutoff capability. Camless or variable valvetiming and lift technologies may be used to shape the post bump to matchthe region where exhaust port pressure is higher than intake portpressure. In addition, these technologies may provide post bump shutoffcapability. Various air systems (shown in FIGS. 5, 6, 7A, 7B, and 11)can be used to overcome lack of airflow (A/F) at low engine speeds witha fixed timing and duration post bump without the shutoff capability.

In one aspect, turbo-chargers may have variable geometry turbines andwaste gates. In another aspect, turbo-chargers may be arranged inseries. In still another aspect, post bump may be shut off at low enginespeeds via, e.g. a lash adjusting mechanism to maintain an acceptableA/F ratio. A conventional turbo-charging scheme (with fixed geometryturbine—see FIG. 8) may also be implemented with shutoff capability.Engine power curves (engine speed and load) may also be manipulated tomaintain acceptable A/F at lower engine speeds if the shutoff capabilityis not available.

In another aspect, the invention provides a combination of in-cylinderEGR and injection rate shaping with a fixed geometry turbo-charger thatmay be optimally matched for low speed engine operation and goodtransient response.

In still another aspect, the invention may be a combination ofin-cylinder EGR and airflow control via an electrically-assistedturbo-charger, or a variable turbine geometry turbo-charger. Includedare injection rate shaping and a turbo-charger that may be optimallymatched for low speed engine operation, in addition to a variableturbine geometry turbo-charger and an electrically assistedturbo-charger.

Injection rate shaping may be provided in combination with the specificstrategy of in-cylinder EGR. The in-cylinder EGR may be accomplished by,e.g opening an exhaust valve during the intake stroke, while the exhaustport pulse pressure is greater than the cylinder and intake portpressure. Injection rate shaping may be a combination of pre-, orpost-combustion injection rate shaping, and changing a shape of the maininjection pulse. Airflow control via, e.g. an electrically assistedturbo-charger, a variable turbine geometry turbo-charger, or a smallerturbine and compressor match may also be included.

In particular, in one embodiment a turbo-charged internal combustioncylinder assembly includes a cylinder having a cylinder head at an endthereof, a combustion chamber with an intake port disposed in thecylinder head, and an intake valve movably disposed in the intake port.The combustion chamber may be communicably connected to theturbo-charger via the intake port so the compressor may providepre-combustion gases to the combustion chamber when the intake valve isopen. An exhaust port is also disposed in the cylinder head, with anexhaust valve movably disposed in the exhaust port that communicablyconnects the combustion chamber to an exhaust manifold. The exhaustvalve may open to exhaust post-combustion gases to the exhaust manifoldwhile the intake valve is substantially closed, and the exhaust valvemay open to admit post-combustion gases to the combustion chamber whilethe intake valve is substantially open. A fuel injector disposed in thecylinder head may admit fuel to the combustion chamber near piston topdead center during, e.g. a compression stroke while both the intake andthe exhaust valves are closed. Such a fuel injector may include a pumpchamber, a fuel-injecting plunger for reciprocating within the pumpchamber, and a discharge nozzle connected to the pump chamber forinjecting fuel into the combustion chamber. A spill valve may bepositioned between the chamber and the nozzle for controlling a rate offuel injection to the combustion chamber, the spill valve having a firstposition providing a maximum fuel injection rate, a second positionproviding a substantially zero fuel injection rate, and at least oneintermediate position providing an intermediate fuel injection ratebetween the maximum fuel injection rate and the zero fuel injectionrate.

In a second embodiment a turbo-charged internal combustion engine systemincludes a cylinder having a combustion chamber with an intake valvedisposed to admit pre-combustion gases to the combustion chamber, and anexhaust port. A first fuel injector may be disposed in the combustionchamber while a second fuel injector disposed in the exhaust port. Anexhaust valve may be disposed to admit post-combustion gases to thecombustion chamber while an exhaust port pressure in the exhaust port ishigher than a combustion chamber pressure in the combustion chamber. Theexhaust valve may be reopened while the exhaust port pressure is higherthan the combustion chamber pressure, and fuel may be injected by afirst fuel injector or a second fuel injector. A purpose for injectingfuel while an exhaust port pressure is higher than a combustion chamberpressure may be to elevate the fuel temperature with the exhaust gas,possibly vaporizing the fuel, and also to mix the fuel with thepre-combustion gases entering through the intake valve. Compression ofthe fuel and air mixture takes place after the intake and the exhaustvalves close, allowing the fuel to autoignite and producing homogenouscharge compression ignition (HCCI) combustion. Autoignition may becontrolled by, e.g. adjusting EGR, the timing of the intake valve event,the compression ratio, or the inlet manifold temperature.

In a third embodiment a fuel injector may be used to control the startof combustion. A first quantity of fuel may be injected into the exhaustport via a second fuel injector while an exhaust port pressure is higherthan a combustion chamber pressure and when an exhaust valve isreopened. A temperature of the first quantity of fuel may be elevated bythe heat of the exhaust gases, and the fuel may be vaporized, when thefirst quantity of fuel mixes with pre-combustion gases entering throughthe intake port. Compression of the fuel and air mixture takes placewhen the intake and exhaust valves have closed. The first quantity offuel, however, is insufficient for auto-ignition to occur. Combustionmay not occur until fuel is also injected by a first fuel injector in aquantity sufficient to auto-ignite. Engine-out emissions may becontrolled by adjusting, e.g. the quantities of fuel injected by thefirst and the second injectors, a timing of fuel injection by the firstfuel injector, the quantity of EGR, a timing of an inlet valve closing,a compression ratio, or an inlet manifold temperature.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

FIG. 1 is a schematic diagram of a turbo-charged internal combustioncylinder assembly according to a first embodiment of the invention;

FIG. 2 is a schematic diagram of a turbo-charged internal combustioncylinder assembly according to the embodiment shown in FIG. 1;

FIG. 3 is a schematic diagram of a turbo-charged internal combustioncylinder assembly according to the embodiment shown in FIG. 1;

FIG. 4 is a graph of a cylinder pressure versus crank angle for aturbo-charged internal combustion engine according to an embodiment ofthe invention;

FIG. 5 is a schematic diagram of a super-charged and turbo-chargedinternal combustion engine according to an embodiment of the invention;

FIG. 6 is a schematic diagram of an air-assisted turbo-charged internalcombustion engine according to an embodiment of the invention;

FIGS. 7A and 7B are schematic diagrams of an electrically-assistedturbo-charged internal combustion engine according to an embodiment ofthe invention;

FIG. 8 is a schematic diagram of a turbo-charged internal combustionengine according to an embodiment of the invention;

FIGS. 9A-9C are schematic diagrams of a fuel injector for use with anembodiment of the invention;

FIG. 10 is a three-quarter view of a turbo-charged internal combustionengine according to an embodiment of the invention;

FIG. 11 is a schematic diagram of a turbo-charged internal combustionengine according to an embodiment of the invention;

FIG. 12 is a three-quarter view of a tractor for use with an embodimentof the invention;

FIG. 13 is a schematic diagram of a turbo-charged internal combustioncylinder assembly according to an embodiment of the invention;

FIGS. 14A-14E show actuators for use with an embodiment of theinvention; and

FIGS. 15A-15E show clutches for use with an embodiment of the invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

In FIG. 1 is shown a turbo-charged internal combustion cylinder assembly100 according to a first embodiment of the invention. Although theembodiments described herein are described in the context of acompression-ignited engine, the concept of the invention could beadapted to other types of ignition as well, such as, e.g. a spark-,hotbulb- or glowplug-ignited engine.

In assembly 100 a cylinder 102 has a cylinder head 104 substantiallyfixedly disposed at one end 105, with a combustion chamber 106 disposedin-cylinder head 104. Cylinder head 104 has an intake port 108 with anintake valve 110 movably disposed in intake port 108, through whichcombustion chamber 106 may communicate with, e.g. a compressor 306.Compressor 306 may, e.g. provide pre-combustion gases 114 through anintake manifold 127, and also through an aftercooler 202 if so equipped,to combustion chamber 106 when intake valve 110 is substantially open.Pre-combustion gases 114 may be, e.g. air, such as a mixture ofnitrogen, oxygen, carbon dioxide, water vapor, and trace elements suchas argon.

As shown in FIG. 2, an exhaust port 128 may also be disposed in-cylinderhead 104, with an exhaust valve 1 16 movably disposed in exhaust port128, through which combustion chamber 106 may communicate with, e.g. anexhaust manifold 126. Post-combustion gases 118 may, e.g exhaust toexhaust manifold 126 and consequently to turbine 307 when exhaust valve116 is substantially open and intake valve 110 is substantially closed.Post-combustion gases 118 may be, e.g. exhaust, such as a mixture ofnitrogen, oxygen, unburned hydrocarbons, carbon dioxide, water vapor,and trace elements such as argon. Although this embodiment of theinvention is described in the context of a four-cycle engine, theprinciple of the invention may be applied to, e.g. a two-stroke engineas well.

As shown in FIG. 3, exhaust valve 116 may, e.g. open and admitpost-combustion gases 118 to combustion chamber 106 while intake valve110 is substantially open. This is an example of internal EGR, which isalso known as ‘post bump’. The bump refers to the exhaust cam lobe. Asshown in FIG. 4, the first time exhaust valve 116 opens, during theexhaust stroke, is the bump. The next time exhaust valve 116 opens,while intake valve 110 is also substantially open, occurs after thebump, or post bump. Post-combustion gases 118 in exhaust port 128 may beat a pressure (710 in FIG. 4) higher than that prevailing (712 in FIG.4) in combustion chamber 106 for at least part of a cycle.

In one embodiment, exhaust valve 116 may be opened with an actuator thatmay be, e.g. a cam, such as a second exhaust cam lobe, a hydraulicactuator, a piezoelectric motor, a voice coil, or a solenoid, as shownin FIG. 14. A second exhaust cam lobe may be formed on the existingexhaust cam or on a separate cam.

Although the various embodiments of the invention are described in thecontext of intake valve 110 and exhaust valve 116 as poppet valves,other types of valves such as, e.g. sleeve valves, rotary valves,louvers, porous membranes, or slide valves could be used as well.

In one embodiment, shown in FIG. 5, pressurized pre-combustion gases 114from compressor 306 may be enhanced by clutch-driven compressor 112.Turbo-charger 302 may, e.g. be comprised of a compressor 306 and aturbine 307 connected by a shaft 305. In this case compressor 306 couldsupply pressurized pre-combustion gases 114 to clutch-driven compressor112 while a clutch 206 is engaged, which will in turn compress furtherthe already pressurized pre-combustion gases 114 and supply them tocombustion chamber 106. This may be particularly useful when, e.g. aspeed or load of an engine is low, reducing the energy available in thepost-combustion gases 118. In this case the amount of energy availablemight be too low to provide adequate power by turbine 307 to drivecompressor 306 through shaft 305. Valve 205 may be closed while, e.g.valve 203 and valve 204 are open.

In one embodiment, pre-combustion gases 114 are supplied solely fromcompressor 306 with clutch 206 disengaged or slipped. This may occurwhen an engine speed or load is high. Under these conditions, energy inpost-combustion gases 118 is adequate to power turbine 307 sufficientlyto drive compressor 306 through shaft 305 and produce adequate boost.Valves 203 and 204 are closed while valve 205 is open.

Clutch 206 is a clutch in the generic sense, meaning any convenientmeans of providing intermittent or interruptable transmission of power,and may be, e.g a single- or a multiple-plate clutch, such as a dry- ora wet-plate clutch, a fluid coupling, such as a hydraulic clutch or ahydrostatic drive, a centrifugal clutch, or an electrostatic clutch.Clutch 206 may also be a continuously variable transmission, such as abelt on pulleys of variable radii, an electric generator driving anelectric motor, or a hydraulic turbine driving an impeller, as shown inFIG. 15.

The pressure prevailing inside a normally-aspirated combustion chamber106 while intake valve 110 is open will, in general, be at or belowsubstantially ambient pressure for at least part of a cycle.Turbo-charged pre-combustion gases 114 will be, in general, at apressure higher than ambient pressure. The pressure prevailing insideexhaust manifold 126 will thus also be higher than that prevailinginside a normally aspirated combustion chamber 106 while intake valve110 is open, and, consequently, while exhaust valve 116 is also open,for at least part of a cycle.

Since the volume of pre-combustion gases 114 delivered to combustionchamber 106 may be controlled by, e.g. engaging and disengaging clutch206, the pressure in the combustion chamber 106 may be controlled and,in particular, raised relative to the pressure in exhaust manifold 126.Since post combustion gases 118 will, in general, move to regions oflower pressure, the volume of post combustion gases 118 returning tocombustion chamber 106 when exhaust valve 116 is open may also becontrolled by adjusting the pressure in combustion chamber 106 relativeto that of exhaust manifold 126.

If, e.g. an engine- or powertrain-controller called for internal exhaustgas recirculation, clutch 206 could be, e.g. disengaged or slipped,reducing the pressure of pressurized pre-combustion gases 114 beingsupplied to combustion chamber 106 relative to that of exhaust manifold126. Since post-combustion gases 118 would then be at a higher pressurethan that of combustion chamber 106, some post-combustion gases 118would be returned to combustion chamber 106 when both intake valve 110and exhaust valve 116 were open.

Since the conditions calling for the addition of post-combustion gases118 may be generally constant over several cycles of engine operation,the control of clutch 206 would not need to change constantly orinstantaneously. Once a particular set operating conditions was reached,the amount of pressure in combustion chamber 106 relative to that ofexhaust manifold 126 could be relatively constant. Post-combustion gases118 could thus, e.g. be used to charge the combustion chamber in themanner of a standing wave.

A take-down at, e.g. the end of a runner of exhaust manifold 126 willpresent a different impedance to an exhaust pulse, i.e. a pulse ofpost-combustion gases 118 than the runner itself. An exhaust pulse maythus travel the length of a runner of exhaust manifold 126, bereflected, or echoed, at the take-down, and return. Such an exhaustpulse would require a certain amount of time to make the round trip. Thetime required would depend on, inter alia the length of the runner andthe pressure differential across the runner.

In one embodiment, the length of the runner could be matched to the timean exhaust pulse would be expected to require to complete the roundtrip. Thus, for a given engine speed, such as e.g. a rated engine speed,or a speed at which the engine was likely to run at highway speeds, thelength of the runner could be set so that, e.g. a previously-emittedexhaust pulse had just enough time to make the round-trip to the end ofthe runner and back just as exhaust valve 116 was opening duringpost-bump. Thus, a returning exhaust pulse would reach exhaust valve 116while it was open to admit post-combustion gases 118 to combustionchamber 106 while intake valve 110 was open, i.e. during post bump.Post-combustion gases 118 could thus, e.g. be used to charge thecombustion chamber in the manner of a standing wave, such as may existin a tuned exhaust. In a similar manner, a length of intake manifold 127runners can be varied to provide lower pressure in combustion chamber106 relative to that of exhaust port 128 which would increase the amountof post-combustion gases 118 transported to the combustion chamber 106when both intake valve 110 and exhaust valve 116 are open. A runnerlength could be varied by, e.g. providing a valve to redirect part ofthe gas flow through a shunt.

The length of a runner however, is not easily changed after an enginehas been built. The timing of the exhaust pulses may thus be tuned foronly a narrow range of engine speeds, and harmonics thereof. In apreferred embodiment, the amount of pressure in combustion chamber 106relative to that of exhaust manifold 126, i.e. the pressure differentialacross the runner, could be adjusted to vary the speed at which a pulsetraveled the length of the runner. The pressure differential could thusbe controlled to substantially ensure that an exhaust pulse wasreturning just as exhaust valve 116 was open during post-bump over awider range of engine speeds. The timing of the opening of exhaust valve116 may be changed to affect the timing of the exhaust pulses as well.

If, e.g. the exhaust pulses were arriving late, the amount of pressurein combustion chamber 106 relative to that of exhaust manifold 126 couldbe lowered. This would result in a larger pressure differential acrossthe exhaust manifold runner, adding impetus to the exhaust pulse,speeding it up so it arrived sooner. The tuning could be accomplished,e.g. by slipping clutch 206, as discussed above, or any method ofpressure control discussed herein, and equivalents thereof.

If, on the other hand, the exhaust pulses were arriving early, theamount of pressure in combustion chamber 106 relative to that of exhaustmanifold 126 could be raised. This would result in a smaller pressuredifferential across the exhaust manifold runner, impeding the exhaustpulse and slowing it down so it arrived later.

In another embodiment, turbo-charger 302 may include, e.g an aftercooler202 to cool pre-combustion gases. In still another embodiment,turbo-charger 302 may include, e.g. a shutoff valve 204 betweencompressor 306 and aftercooler 202 to reduce a flow of pre-combustiongases. Shutoff valve 204 may be used to control the pressure in thecombustion chamber 106 relative to the pressure in e.g. exhaust manifold126 in the same manner as the clutch-driven compressor. Shutoff valve204 may be, e.g. controlled or actuated electrically, hydraulically,pneumatically, or mechanically.

If, e.g. an engine or powertrain controller called for internal exhaustgas recirculation, shutoff valve 204 could be, e.g. partially orcompletely closed, reducing the pressure of pressurized pre-combustiongases 114 being supplied to combustion chamber 106 relative to that ofexhaust manifold 126. Since post-combustion gases 118 would then be at ahigher pressure than that of combustion chamber 106, somepost-combustion gases 118 would be returned to combustion chamber 106when both intake valve 110 and exhaust valve 116 were open.

In one embodiment, shown in FIG. 6, turbo-charger 302 may be, e.g.comprised of a compressor 306 and a turbine 304 connected by a shaft305. Turbo-charger 302 may be, e.g. a fixed-geometry turbo-charger, avariable-geometry turbo-charger, or a low inertia turbo-charger. Inanother embodiment, turbo-charger 302 may include, e.g. a source ofcompressed air 304 to supply pre-combustion gases to a compressor 306 ofturbo-charger 302. In the alternative, source of compressed air 304 maysupply pre-combustion gases 114 to combustion chamber 106 directly.

Source of compressed air 304 may be used to supply pre-combustion gasesto a compressor 306 if, e.g. there are not sufficient post-combustiongases 118 to drive a compressor 306 adequately. This may be the caseunder operating conditions such as, e.g. start-up or low-speedoperation. Source of compressed air 304 may also be used to control thepressure in the combustion chamber 106 relative to the pressure in e.g.exhaust manifold 126 in the same manner as the clutch-driven compressor.

If, e.g. an engine or powertrain controller called for internal exhaustgas recirculation, source of compressed air 304 could be, e.g. partiallyor completely restricted, reducing the pressure of pressurizedpre-combustion gases 114 being supplied to combustion chamber 106relative to that of exhaust manifold 126. Since post-combustion gases118 would then be at a higher pressure than that of combustion chamber106, some post-combustion gases 118 would be returned to combustionchamber 106 when both intake valve 110 and exhaust valve 116 were open.

In one embodiment, turbo-charger 302 may include, e.g. a shutoff valve308 between source of compressed air 304 and compressor 306. In analternative embodiment, source of compressed air 304 may be rechargedbetween uses.

Engine performance may also be enhanced by electrically assistingturbo-charger 302 for better airflow control. An electrically assistedturbo-charger may improve low engine speed operation by providingcombustion air to the engine such as, e.g. with a long duration ofexhaust valve reopening.

In one embodiment, shown in FIG. 7, turbo-charger 302 may include, e.g.an electric motor 402 to assist a compressor 306 of turbo-charger 302.Electric motor 402 may be used to turn compressor 306 if, e.g. there arenot sufficient post-combustion gases 118 to drive compressor 306adequately. This may be the case under operating conditions such as,e.g. start-up or low-speed operation. Electric motor 402 may be used tocontrol the pressure in the combustion chamber 106 relative to thepressure in e.g. exhaust manifold 126 in the same manner as theclutch-driven compressor.

If, e.g. an engine or powertrain controller called for internal exhaustgas recirculation, a speed or a torque of electric motor 402 could be,e.g. partially or completely reduced, reducing the pressure ofpressurized pre-combustion gases 114 being supplied to combustionchamber 106 relative to that of exhaust manifold 126. Sincepost-combustion gases 118 would then be at a higher pressure than thatof combustion chamber 106, some post-combustion gases 118 would bereturned to combustion chamber 106 when both intake valve 110 andexhaust valve 116 were open.

Engine performance may also be enhanced by including a variable geometryturbo-charger. A variable geometry turbo-charger improves low enginespeed operation. Turbo-charger 302 may be, e.g. a fixed-geometryturbo-charger, a variable-geometry turbo-charger, or a low inertiaturbo-charger. In one embodiment, also shown in FIG. 8, turbo-charger302 may include, e.g. an 87 mm diameter compressor and a 76 mm diameterturbine.

In another embodiment, also shown in FIG. 8, turbo-charger 302 mayinclude, e.g. a waste gate 404. Waste gate 404 may be used to controlthe pressure in the combustion chamber 106 relative to the pressure ine.g. exhaust manifold 126 in a manner similar to the clutch-drivencompressor. Having wastegate 404 as an option allows the use of asmaller turbine housing or a smaller turbine wheel and may also allowhigher airflow at lower engine speeds and torques. Use of a smallerturbine housing or a smaller turbine wheel may further result inimproved turbo-charger response during transient engine conditions. Adrawback of using a smaller turbine housing or a smaller turbine wheel,on the other hand, may be excessive turbine boost and speed at higherengine speeds and torque. Higher turbine boost may produce highercylinder pressures and increase thermal loading of the aftercooler, aswell.

If, e.g an engine or powertrain controller called for lower cylinderpressures, waste gate 404 could be, e.g. partially or completely opened,reducing the pressure of pressurized pre-combustion gases 114 beingsupplied to combustion chamber 106. Opening the wastegate 404 may alsoreduce turbine speed, consequently prolonging the lives of thecompressor and the turbine and reducing thermal loading of aftercooler202.

As shown in FIG. 1, a fuel injector 120 may be disposed in-cylinder head104 to admit fuel 122 to combustion chamber 106. As shown in FIG. 9,fuel injector 120 may include a pump chamber 502, a fuel-injectingplunger 504 that reciprocates within pump chamber 502, and a dischargenozzle 506 connected to pump chamber 502 for injecting fuel 122 intocombustion chamber 106.

A spill valve 508 may be positioned between pump chamber 502 anddischarge nozzle 506 for controlling a rate of fuel injection tocombustion chamber 106. The spill valve 508 shown in FIG. 9 has threepositions. Spill valve 508 has a first position 514 providing a maximumfuel injection rate when spill valve plunger 518 is closed (FIG. 9C), asecond position 510 providing a substantially zero fuel injection ratewhen the spill valve plunger 518 is fully open (FIG. 9A), and at leastone intermediate position 512 providing an intermediate fuel injectionrate between maximum fuel injection rate and zero fuel injection ratewhen the spill valve plunger 518 is partially closed (FIG. 9B).

In operation, fuel 122 may be fed to the fuel injector 120 by the fuelsupply line 516. The fuel-injecting plunger 504 pressurizes the fuel122, and the spill valve 508 controls the spilling of fuel above thefuel-injecting plunger 504. When the spill valve 508 is completely open,fuel is spilled at a rapid rate, and no increase in the fuel pressureoccurs. When the spill valve 508 is partially closed, the fuel above thefuel-injecting plunger 504 may be pressurized, but due to the slightspilling action the spilling effectively reduces the cam velocity. Whenthe spill valve 508 is completely closed, the fuel may be completelypressurized and the discharge nozzle 506 opens.

This spilling action may be electronically controlled, and may occur,for example, during the first (and critical) five to ten crank degreesof fuel injection. This may be especially important for urban operation.It should be appreciated, however, that the electronically controlledspilling action may be performed at any time, and it is not strictlyconfined to the first five to ten crank degrees of fuel injection.

There may be an opportunity to reduce cam velocity associated withinjection rate shaping. The effective reduction in cam velocity may bedependent on the spill area offered by the configuration of the spillvalve 508. The duration of the spilling action may be dependent on thereaction capability of the spill valve 508 (i.e., how quickly the valvemay be opened or closed). In a preferred embodiment, the three-positionspill valve 508 should be capable of moving to the partially closedposition and dwelling at this position for approximately one millisecondbefore completely closing.

In one embodiment, spill valve 508 actuation may be, e.g. controlledelectronically, and can open at any time in an engine cycle. In oneembodiment, spill valve 508 may be, e.g. actuated by a solenoid. Inanother embodiment, spill valve 508 may be, e.g. a magnetic-latchingspill valve. In one embodiment, spill valve 508 may be, e.g. capable ofdwelling at an intermediate position 514 for about one millisecond. Inanother embodiment, spill valve 508 may be, e.g. capable of attaining atleast one intermediate position 514 during a first five to ten degreesof crankshaft rotation.

In one embodiment, an intermediate fuel injection rate may be, e.g. usedwhen a load on the engine is between 20% of a maximum load and 100% of amaximum load.

Injection rate shaping may occur, e.g. from peak torque to rated enginespeed. This will allow further reduction in NOx in addition to NOxreduction due to in-cylinder EGR.

An added benefit of injection rate shaping may also be higher stacktemperatures, which will improve after treatment efficiencies. Additionof injection rate shaping may also allow a reduction in the duration ofexhaust valve reopening which helps in improving the low engine speedoperation while maintaining NOx at an acceptable level.

In another embodiment, an intermediate fuel injection rate may be, e.g.used when a speed of the engine is between the speed at which a peaktorque occurs and a rated engine speed.

In another embodiment, as shown in FIG. 4, first and second exhaust camlobes 702, 704 may, e.g. be operably disposed to open exhaust valve 116.During an exhaust stroke, first cam lobe 702 may, e.g. open exhaustvalve 116 to exhaust post-combustion gases 118 to exhaust manifold 126while intake valve 110 is substantially closed. Then, during thefollowing intake stroke, second cam lobe 704 may open exhaust valve 116to admit post-combustion gases 118 to combustion chamber 108 whileintake valve 110 is substantially open and exhaust port pressure 710 ismomentarily higher than cylinder pressure 712.

In one embodiment, a maximum fuel injection rate may be used whenexhaust valve 116 is substantially closed. In another embodiment, anintermediate fuel injection rate may be used when exhaust valve 116 isopen to admit post-combustion gases to combustion chamber 106.

In FIG. 10 is shown a turbo-charged internal combustion engine 1000according to a second embodiment of the invention. In FIG. 10 aplurality of cylinder assemblies 100 may be combined to form an engine1000. In this embodiment, six cylinder assemblies 100 are combined toform engine 1000. In an alternative embodiment, four cylinder assemblies100 are combined to form engine 1000. In a further alternativeembodiment, eight cylinder assemblies 100 are combined to form engine1000. Various numbers of cylinder assemblies 100 may be arranged in,e.g. an in-line, a vee, a radial, an opposed, or a flat configurationwithout departing from the spirit of the invention.

As shown in FIG. 11, six cylinder assemblies 100 are connected to afirst and second turbo-chargers 804, 806. In this embodiment, half ofthe plurality of cylinder assemblies 100 communicates substantiallyexclusively via exhaust manifold 126 with first turbo-charger 804, whilethe other half communicates substantially exclusively via exhaustmanifold 126 with second turbo-charger 806. For example, a plurality ofsix cylinder assemblies 100 would include a first three cylinders 908and a second three cylinders 910, while a plurality of four cylinderassemblies 100 would include a first two cylinders and a second twocylinders.

If, e.g. a firing order of an in-line six-cylinder engine were1-5-3-6-2-4, exhaust pulses from the first, second, and third cylinderscould be applied to first turbo-charger 804, while exhaust pulses fromthe fourth, fifth, and sixth cylinders could be applied to secondturbo-charger 806. This would distribute the exhaust pulses seen by eachof turbo-chargers 804, 806 substantially 240° apart, rather than 120° aswould be the case if all of the runners met at a single turbo-charger.

Exhaust pulses arriving 240° apart may be less likely to be diluted bypreceding or following pulses than if they were arriving 120° apart.There may thus be less interpulse interference, to paraphrasecommunications jargon. Exhaust pulses arriving 240° apart may thus beeasier to measure and control than those spaced more closely together.Furthermore, two turbo-chargers 804, 806 could be smaller than a singleturbo-charger, and thus of lower inertia. Two turbo-chargers 804, 806 oflower inertia would be easier to wind up. Two turbo-chargers 804, 806may thus be able to respond to transients faster than a singleturbo-charger.

If the engine were equipped with an aftercooler 202, the output fromturbo-chargers 804, 806 could be plumbed separately through aftercooler202, to substantially preserve the independence of the outputs.

More than two turbo-chargers could be used, without departing from thespirit of the invention. At the limit, an individual turbo-charger couldbe used for each cylinder. In that case the pressure drop across eachrunner could be controlled individually, resulting in relatively preciseinternal exhaust gas recirculation.

In FIG. 12 is shown a third embodiment of the invention. In FIG. 12 aturbo-charged internal combustion engine 800 may be installed in atractor 1200. An embodiment of the invention could also be used in, e.g.stationary applications, marine applications, agricultural equipment,earth moving equipment, locomotives, or aircraft, includinglighter-than-air craft.

In FIG. 13 is shown a turbo-charged internal combustion engine system1200 according to a fourth embodiment of the invention. System 1200includes a cylinder 1202 having a combustion chamber 1204 with an intakeport 1224 and an exhaust port 1206. An intake valve 1212 may be disposedin intake port 1224 to admit pre-combustion gases to combustion chamber1204, while an exhaust valve 1214 may be disposed in exhaust port 1206to exhaust post-combustion gases 1220 to an exhaust manifold 1222.

Exhaust valve 1214 may be operable to admit post-combustion gases 1220to combustion chamber 1204 while intake valve 1212 is open and anexhaust port pressure 1216 in exhaust port 1206 is higher than acombustion chamber pressure 1218 in combustion chamber 1204.

A first fuel injector 1208 may be disposed, e.g. in combustion chamber1204, or, in the alternative, in intake port 1224. This may be the casein, e.g. a spark-ignited engine. A second fuel injector 1210 may bedisposed, e.g. in exhaust port 1206. Each of first and second fuelinjectors 1208, 1210 may be equipped individually for injection rate orpulse shaping.

First fuel injector 1208 may admit fuel to combustion chamber 1204 whileintake 1212 is open, e.g. during an intake stroke. First fuel injector1208 may also admit fuel to combustion chamber 1204 while intake andexhaust valves 1212, 1214 are closed, i.e. near the end of a compressionstroke. This may be the case in, e.g. a compression ignition engine.

In one embodiment, both first and second fuel injectors 1208, 1210 mayadmit fuel to combustion chamber 1204 while intake and exhaust valves1212, 1214 are open and exhaust port pressure 1216 is higher thancombustion chamber pressure 1218, i.e. during internal exhaust gasrecirculation. In another embodiment, only second fuel injector 1210admits fuel to combustion chamber 1204 during internal EGR.

Since post-combustion gases 1220 are comprised largely of, e.g. carbondioxide and water, both of which are common fire retardants, combustionis not likely to occur prematurely. That is, the fuel being injectedinto the stream of post-combustion gases 1220 being returned tocombustion chamber 1204 will not ignite in the absence of oxygen. Thefuel will, however, be warmed by the heat of post-combustion gases 1220,in the manner of a regenerator.

Raising the temperature of incoming fuel will increase the likelihoodthat when combustion does occur it will be substantially complete, thusreducing emission of unburned hydrocarbons. Also, raising thetemperature at which heat is added to the combustion chamber by usingheat that would otherwise be rejected may result in efficiencies over anengine in which fuel is injected cold.

In a fifth embodiment, the invention includes a method of controllingcombustion in an internal combustion engine comprising the steps oftransferring a first quantity of fuel and first post-combustion gases toa combustion chamber through an exhaust port to raise a temperature ofsaid first quantity of fuel while an exhaust port pressure in theexhaust port is higher than a first combustion chamber pressure.

The A/F ratio and the temperature will be too low to ignite the fuelduring this time, but the post-combustion gases will have transferredtheir heat to the fuel, warming it up. Pre-combustion gases are thentransferred into the combustion chamber at the first combustion chamberpressure, mixing with the first quantity of fuel, the pre-combustiongases, and the first post-combustion gases to form a substantiallyhomogenous mixture. The intake and exhaust ports are substantiallysealed, and heat is added to the combustion chamber by raising the firstcombustion chamber pressure to a second combustion chamber pressuresubstantially higher than the first combustion chamber pressure, i.e.during a compression stroke. Fuel injected by fuel injector 1208disposed in combustion chamber 1204 may, e.g. initiate combustion whenthe piston is near top dead center during a compression stroke.

In a sixth embodiment, the step of transferring a first quantity of fuelinto combustion chamber occurs substantially between a crankshaft angleof 380° and a crankshaft angle of 470°, i.e. during internal EGR.

In a seventh embodiment, the step of transferring the first quantity offuel and first post-combustion gases to the combustion chamber throughthe exhaust port comprises further the step of injecting the firstquantity of fuel into the exhaust port.

In a ninth embodiment, the step of injecting a second quantity of fuelinto the combustion chamber occurs substantially between a crankshaftangle of 380° and a crankshaft angle of 470°.

In an eleventh embodiment, the step of injecting both a first quantityof fuel and a second quantity of fuel into the combustion chamber occurssubstantially between a crankshaft angle of 380° and a crankshaft angleof 470°.

In a twelfth embodiment, the step of injecting a third quantity of fuelinto the combustion chamber occurs substantially between the angles of−25° to 5° after top dead center (ATDC).

In a thirteenth embodiment, the step of injecting a third quantity offuel further comprises the step of shaping a rate of fuel injection.

In a fourteenth embodiment, the step of mixing 1) a first quantity offuel, pre-combustion gases, and first post-combustion gases, 2) a secondquantity of fuel, pre-combustion gases, and first post-combustion gases,and 3) first and second quantities of fuel, pre-combustion gases, andfirst post-combustion gases to form a substantially homogenous mixtureoccurs substantially between a crankshaft angle of 400° and a crankshaftangle of 695°, i.e. during an intake stroke.

Injection rate shaping capability with air systems improvement (i.e.variable geometry turbo-charger, electrically assisted turbo-charger,light inertia turbo-charger (ceramic turbine wheel, titanium aluminideturbine wheel, smaller compressor and turbine trim)), may improve lowengine speed operation, and also help with transient engine response.

Steady state and transient emissions and engine performance data showsthat in-cylinder EGR in conjunction with rate shaping with properlymatched turbo-charger can result in significant reduction in NOx(reduction in excess of 20% NOx) with minimal increase in particulates(PM). In one example, overall airflow dropped about 10% due toin-cylinder EGR. As a result, a much smaller turbo-charger may be used(which helps in engine response) than would be used on a typical12-liter engine. Also rate shaping enabled enhanced turbo-chargerresponse by providing hotter pre-turbine temperatures.

The lower airflow through the engine (due to in-cylinder EGR) may permitusing a lower volume exhaust manifold. This also helps on transientengine response, which lowers particulates, and on maintaining higherexhaust port pressure, thus allowing more in-cylinder EGR, which in turnhelps to reduce NOx.

In Table 1 is shown representative engine test data: TABLE 1 OICA*Transient* OICA** Transient** NOx (g/bhp*hr) 3.245 3.210 2.435 2.450  PM(g/bhp*hr) 0.048 0.075 0.060 0.093*Injection Rate Shaping Alone**Injection Rate Shaping + In-cylinder Exhaust Gas Recirculation

Although the preferred embodiment above discloses the use of asolenoid-type valve, it is contemplated that a magnetic latching valvemay optionally be used. In addition, although a three-position spillvalve is disclosed in the preferred embodiment, alternatively a spillvalve may be used having more than three positions in order to providean even more finely controlled flow of fuel.

While the invention has been described in detail above, the invention isnot intended to be limited to the specific embodiments as described. Itis evident that those skilled in the art may now make numerous uses andmodifications of and departures from the specific embodiments describedherein without departing from the inventive concepts.

1. A method of controlling combustion in an internal combustion enginecomprising the steps of: transferring a first quantity of fuel and firstpost-combustion gases to a combustion chamber through an exhaust port toraise a temperature of said first quantity of fuel while an exhaust portpressure in said exhaust port is higher than a first combustion chamberpressure; transferring pre-combustion gases into said combustion chamberat said first combustion chamber pressure; mixing said first quantity offuel, said pre-combustion gases, and said first post-combustion gases toform a substantially homogenous mixture; substantially sealing saidexhaust port from said combustion chamber; adding heat to saidcombustion chamber by raising said first combustion chamber pressure toa second combustion chamber pressure substantially higher than saidfirst combustion chamber pressure; transferring a second quantity offuel to said combustion chamber; substantially combusting said first andsecond quantities of fuel with said heat to form second post-combustiongases; exhausting said second post-combustion gases through said exhaustport.
 2. The method of controlling combustion in an internal combustionengine of claim 1, wherein the step of transferring said first quantityof fuel and first post-combustion gases to said combustion chambercomprises further the step of injecting said first quantity of fuel intosaid exhaust port or exhaust manifold.
 3. The method of controllingcombustion in an internal combustion engine of claim 1, wherein the stepof injecting said second quantity of fuel further comprises the step ofshaping a rate of said fuel injection.
 4. The method of controllingcombustion in an internal combustion engine of claim 1, wherein the stepof substantially combusting said first and second quantities of fuelwith said heat to form second post-combustion gases comprises furtherthe steps of: injecting a third quantity of fuel into said combustionchamber; substantially combusting said third quantity of fuel with saidheat.
 5. The method of controlling combustion in an internal combustionengine of claim 4, wherein the step of injecting said third quantity offuel comprises further the step of shaping a rate of said fuelinjection.